JP2004124953A - Ball bearing - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a ball bearing preventing seizure by reducing heat generation in accompany with high-speed rotation, elongating the rolling fatigue life, and reducing the cost. <P>SOLUTION: A curvature radius of a cross section in the width direction of an inner raceway groove 15 is set to 0.535Da-0.545Da with respect to a diameter Da of a ball 13, a curvature radius of a cross section in the width direction of an outer raceway groove 15 is set to 0.515Da-0.545Da with respect to the diameter Da of the ball 13, a contact angle of the ball 13 to the inner raceway groove 15 and the outer raceway groove 16 is set to 17°-26°, and the diameter of the ball 13 is set to 0.4H-0.56H with respect to a bearing cross-sectional height H (1/2 of value obtained by subtracting inner diameter of inner ring 11 from outer diameter of outer ring 12). <P>COPYRIGHT: (C)2004,JPO

Description

【００２７】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）のアキシアルばね定数を図１に示す。同図に示すように、番号１〜２５の中で回転時のアキシアルばね定数が最も高い玉軸受は番号１の玉軸受であり、逆に回転時のアキシアルばね定数が最も低い玉軸受は番号２５の玉軸受である。
番号１の玉軸受と番号２５の玉軸受とを比較すると、番号１の玉軸受は玉の直径をＤａとすると内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが０．５２Ｄａであるのに対し、番号２５の玉軸受は内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが０．６０Ｄａであることから、内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが大きいほど回転時のアキシアルばね定数が低下し、回転時のアキシアル剛性が低くなることがわかる。
【００２８】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の内輪ＰＶ値を図２に示す。同図に示すように、番号１〜２５の中で回転時の内輪ＰＶ値が最も高い玉軸受は番号１の玉軸受であり、逆に回転時の内輪ＰＶ値が最も低い玉軸受は番号２０及び２５の玉軸受であり、番号５、１０及び１５の玉軸受も同様である。
【００２９】
番号５、１０、１５、２０及び２５の玉軸受は玉の直径をＤａとすると内側軌道溝の幅方向断面の曲率半径Ｒｉが０．６０Ｄａであり、このことから、外側軌道溝の幅方向断面の曲率半径Ｒｏが変わっても内側軌道溝の幅方向断面の曲率半径Ｒｉが同じ値であれば内輪ＰＶ値は変化しないことがわかり、さらに内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど内輪ＰＶ値を低く抑えられることがわかる。
【００３０】
表１の番号１〜２５で示される各玉軸受の内輪を２５０００ｍｉｎ^−１で回転させた時の内輪ＰＶ値と外輪ＰＶ値を図３に示す。同図に示すように、玉軸受の内輪を２５０００ｍｉｎ^−１の高速で回転させた場合には、外輪ＰＶ値よりも内輪ＰＶ値のほうが大きくなる。これは、高速回転になると外輪コントロール（外輪と玉とが純転がりに近い状態）になり、玉の自転軸が外輪側の接触角を基準になるため、内輪側の滑りが大きくなり、ＰＶ値が増大するからである。したがって、焼付きを判断するための指標としては、内輪ＰＶ値を採用することが望ましい。
【００３１】
表１の番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の発熱量を図４に示す。同図に示すように、番号１〜２５の中で発熱量が最も高い玉軸受は番号１の玉軸受であり、逆に発熱量が最も低い玉軸受は番号１０，１５，２０，２５の玉軸受である。番号１０，１５，２０，２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａであり、このことから、内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど高速回転時の発熱量が低下し、焼付きを抑制できることがわかる。
【００３２】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の転がり疲労寿命を図５に示す。同図に示すように、番号１〜２５の中で転がり疲労寿命が最も長い玉軸受は番号６の玉軸受であり、逆に転がり疲労寿命が最も短い玉軸受は番号２５の玉軸受である。
番号６の玉軸受と番号２５の玉軸受とを比較すると、番号６の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．５２Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．５４Ｄａであるのに対し、番号２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．６０Ｄａであることから、内側軌道溝および外側軌道溝の幅方向断面の曲率半径Ｒｉ，Ｒｏが大きいほど転がり疲労寿命が低下することがわかる。
【００３３】
表１に番号１〜２５で示される各玉軸受の回転時（２５０００ｍｉｎ^−１）の静的許容アキシアル荷重を図６に示す。同図に示すように、番号１〜２５の中で静的許容アキシアル荷重が最も高い玉軸受は番号１の玉軸受であり、逆に静的許容アキシアル荷重が最も低い玉軸受は番号５，１０，１５，２０，２５の玉軸受である。
【００３４】
番号５，１０，１５，２０，２５の玉軸受は内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．６０Ｄａであり、このことから、内側軌道溝の幅方向断面の曲率半径Ｒｉが大きいほど静的許容アキシアル荷重が低下してしまうことがわかる。
したがって、玉軸受の発熱量やＰＶ値を低減するためには、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を大きくすれば良いが、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を大きくすると軸受の剛性や寿命が低下すると共に静的許容アキシアル荷重が低下してしまう。これと反対に、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を小さくすると発熱量やＰＶ値が上昇してしまう。
【００３５】
工作機械主軸支持用としては、焼付き指標、発熱量、アキシアル剛性、転がり疲労寿命、静的許容アキシアル荷重の全てにおいてバランスの良好な軸受が求められており、このような観点から、本発明者らは、番号１〜２５で示される各玉軸受のアキシアルばね定数、内輪ＰＶ値、発熱量、転がり疲労寿命、静的許容アキシアル荷重の計算結果を絶対値評価で比較した。ここでの絶対値評価は、各条件において最も優れる値を１００、最も劣る値を０とし、その差分を指数化して絶対値評価を行った。また、高速条件下で最も重要となる焼付き指標−内輪ＰＶ値は重要度を上げ、ランクを２倍として計算した。
【００３６】
上記の絶対値評価で使用した計算式を下記に示す。
（１）剛性（Ｋａ）
ＫａＲａｎｋ＝１００−（Ｋａ_Ｍａｘ−Ｋａ）／（（Ｋａ_Ｍａｘ−Ｋａ_Ｍｉｎ）／１００）
Ｋａ_Ｍａｘ；図１の中で最大値を示すアキシアルばね定数
Ｋａ_Ｍｉｎ；図１の中で最小値を示すアキシアルばね定数
（２）内輪ＰＶ値（Ｐｖ）
ＰｖＲａｎｋ＝（Ｐｖ_Ｍａｘ−Ｐｖ）／（（Ｐｖ_Ｍａｘ−Ｐｖ_Ｍｉｎ）／１００）
Ｐｖ_Ｍａｘ；図２の中で最大値を示す内輪ＰＶ値
Ｐｖ_Ｍｉｎ；図２の中で最小値を示す内輪ＰＶ値
（３）発熱量（Ｗ）
ＷＲａｎｋ＝（Ｗ_Ｍａｘ−Ｗ）／（（Ｗ_Ｍａｘ−Ｗ_Ｍｉｎ）／１００）
Ｗ_Ｍａｘ；図４の中で最大値を示す発熱量
Ｗ_Ｍｉｎ；図４の中で最小値を示す発熱量
（４）転がり疲れ寿命（Ｌ）
ＬＲａｎｋ＝１００−（Ｌ_Ｍａｘ−Ｌ）／（（Ｌ_Ｍａｘ−Ｌ_Ｍｉｎ）／１００）
Ｌ_Ｍａｘ；図５の中で最大値を示す転がり疲れ寿命
Ｌ_Ｍｉｎ；図５の中で最小値を示す転がり疲れ寿命
（５）静的許容アキシアル荷重（Ｆａ）
ＦａＲａｎｋ＝１００−（Ｆａ_Ｍａｘ−Ｆａ）／（（Ｆａ_Ｍａｘ−Ｆａ_Ｍｉｎ）／１００）
Ｆａ_Ｍａｘ；図６の中で最大値を示す静的許容アキシアル荷重
Ｆａ_Ｍｉｎ；図６の中で最小値を示す静的許容アキシアル荷重
（６）ＴＯＴＡＬ−Ｒａｎｋ
ＫａＲａｎｋ＋（ＰｖＲａｎｋ×２）＋ＷＲａｎｋ＋ＬＲａｎｋ＋ＦａＲａｎｋ
上述したＫａＲａｎｋ、ＰｖＲａｎｋ、ＷＲａｎｋ、ＬＲａｎｋ、ＦａＲａｎｋ、ＴＯＴＡＬ−Ｒａｎｋの計算結果を表２に示す。
【００３７】
【表２】

【００３８】
ＴＯＴＡＬ−Ｒａｎｋの評価点は、番号２及び７の玉軸受が最も優れており、最適な仕様と言える。番号２及び７のＴＯＴＡＬ−Ｒａｎｋは、ほぼ同じ値である。
表２のＲａｎｋ係数からいずれかの指標が１００点を取るような軸受仕様は、いずれかの値が０もしくは０に近い値になり、両極端となる。このため、工作機械主軸用としては、あまり良い仕様とは言えない。
【００３９】
表２に示される計算結果から、工作機械主軸支持用として最も適している玉軸受は玉の直径をＤａとすると内側軌道溝の幅方向断面の曲率半径ＲｉがＲｉ＝０．５４Ｄａであり、また外側軌道溝の幅方向断面の曲率半径ＲｏがＲｏ＝０．５２Ｄａもしくは０．５４Ｄａであることがわかる。しかし、軸受を実際に製作する際には製作誤差も含まれるため、内側軌道溝の幅方向断面の曲率半径ＲｉをＲｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、外側軌道溝の幅方向断面の曲率半径ＲｏをＲｏ＝０．５１５Ｄａ〜０．５４５Ｄａの範囲とするのが望ましい。
【００４０】
次に、本発明者らは、玉の接触角の影響を確認するために、内輪内径を７０ｍｍ、外輪外径を１１０ｍｍ、玉径をＤａ＝８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ＝Ｒｏ＝０．５４Ｄａ、内外輪材質をＳＵＪ２、玉材質をＳｉ_３Ｎ_４とし、玉の接触角をα＝１２°、１５°、１８°、２０°、２２°、２５°、３０°とした場合の内輪ＰＶ値、外輪ＰＶ値、発熱量について計算した。なお、ここでの計算は、回転数；２５０００ｍｉｎ^−１、ＤＢ組合せ（間座幅２０ｍｍ）、軸嵌め合い；Ｔ２１（しまり嵌め、嵌め合いタイト量；２１μｍ）、ハウジング嵌め合い；Ｌ８（すきま嵌め、嵌め合いルーズ量；８μｍ）、潤滑；オイルエア潤滑、組込み時予圧荷重；アキシアルばね定数１００Ｎ／μｍになる予圧荷重（この組込み時のアキシアルばね定数１００Ｎ／μｍは切削加工可能な剛性目安として使用）、温度条件；外輪温度上昇２４℃、内外輪温度差１６℃の条件で計算した。その計算結果を図７乃至図９に示す。
【００４１】
図７に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなるため、内輪ＰＶ値が急激に上昇する。したがって、内輪ＰＶ値を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を３０°より小さい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角を例えば１２°にするとアキシアル剛性が小さくなり、組込み時のアキシアルばね定数を１００Ｎに設定しようとした場合に過大な予圧荷重をかけなくてはならず、回転時の予圧も大きくなる。このため、図７に示すように、内側軌道溝および外側軌道溝に対する玉の接触角を１５°とした玉軸受よりも内輪ＰＶ値が増大するため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４Ｄａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１５°〜２５°とすることが望ましい。
【００４２】
また、図８に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなるため、発熱量が急激に上昇する。したがって、発熱量を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を３０°より小さい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角が１５°の場合には、内側軌道溝および外側軌道溝に対する玉の接触角が１８°の場合よりも発熱量が増大する。このため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４Ｄａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１８°〜２５°とすることが望ましい。
【００４３】
さらに、図９に示すように、内側軌道溝および外側軌道溝に対する玉の接触角がα＝１５°以下の場合には、上記と同様の理由により、外輪ＰＶ値が増大する。したがって、外輪ＰＶ値を低く抑えるためには、内側軌道溝および外側軌道溝に対する玉の接触角を１５°より大きい角度にする必要があるが、内側軌道溝および外側軌道溝に対する玉の接触角が３０°の場合には、ジャイロモーメントの影響により玉の滑りが大きくなり、内側軌道溝および外側軌道溝に対する玉の接触角が２５°の場合よりも外輪ＰＶ値が上昇する。このため、内側軌道溝および外側軌道溝の幅方向断面の曲率半径を表１に示す番号７のようにＲｉ，Ｒｏ＝０．５４ａと設定した場合には、内側軌道溝および外側軌道溝に対する玉の接触角を１８°〜２５°とすることが望ましい。
【００４４】
以上、図７乃至図９の結果より内側軌道溝及び外側軌道溝に対する玉の接触角は１８°〜２５°とすることが望ましいと言えるが、軸受を実際に製作する際には製作誤差も含まれるため、内側軌道溝及び外側軌道溝に対する玉の接触角は１７°〜２６°とすることが望ましい。
次に、本発明者らは、玉径の確認するために、内輪内径を７０ｍｍ、外輪外径を１１０ｍｍ、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ＝Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＵＪ２、玉材質をＳｉ_３Ｎ_４とし、玉径、玉数および断面比（＝玉径／（外輪外径−内輪内径）／２）×１００）を表３に示す組合せとした場合の内輪ＰＶ値を計算した。なお、ここでの計算は、回転数；２５０００ｍｉｎ^−１、ＤＢ組合せ（間座幅２０ｍｍ）、軸嵌め合い；Ｔ２１（しまり嵌め、嵌め合いタイト量；２１μｍ）、ハウジング嵌め合い；Ｌ８（すきま嵌め、嵌め合いルーズ量；８μｍ）、潤滑；オイルエア潤滑、組込み時予圧荷重；アキシアルばね定数１００Ｎ／μｍになる予圧荷重（この組込み時のアキシアルばね定数１００Ｎ／μｍは切削加工可能な剛性目安として使用）、温度条件；外輪温度上昇２４℃の条件で計算した。その計算結果を図１０に示す。
【００４５】
【表３】

【００４６】
図１０から明らかなように、玉の直径を７．１４４ｍｍ、７．９３８ｍｍとした場合には、内輪と外輪との温度差が大きくなると内輪ＰＶ値が急激に上昇し、内輪と外輪との温度差が３０°付近では内輪ＰＶ値が１９００〜２１００ＭＰａ・ｍ／ｓ程度となる。また、玉の直径を１１．９０６ｍｍ、１２．７ｍｍとした場合は、内輪ＰＶ値の変化は緩やかであるが、元々のＰＶ値自体が高く、内輪と外輪の温度差が３０°付近は、前述の７．１４４ｍｍ、７．９３８ｍｍと同じく内輪ＰＶ値が１９００〜２１００ＭＰａ・ｍ／ｓ程度となる。これに対し、玉の直径を８．７３１ｍｍ、９．５２５ｍｍ、１０．３１９ｍｍ、１１．１１２ｍｍとした場合には、内輪と外輪との温度差が大きくなると内輪ＰＶ値も上昇するが、内輪と外輪との温度差が３０°付近では内輪ＰＶ値が１８００ＭＰａ・ｍ／ｓ程度となる。したがって、軸受断面高さ（外輪の外径から内輪の内径を差し引いた値の１／２）をＨとすると玉の直径ＤａがＤａ＝０．４３７Ｈ〜０．５５６Ｈである場合には、内輪と外輪との温度差に対する内輪ＰＶ値の変化率がほぼ同等となるので、玉の直径ＤａをＤａ＝０．４３７Ｈ〜０．５５６Ｈに設定することにより、内輪と外輪との間に温度差があってもＰＶ値を低くすることができる。
【００４７】
図１１は、本発明の一実施形態に係る玉軸受の部分断面図である。同図に示すように、本発明の一実施形態に係る玉軸受１０は、リング状に形成された内輪１１と、この内輪１１の外周に同心円状に配置された外輪１２とを備えており、内輪１１及び外輪１２のうち少なくとも一方は、重量比でＣを０．２〜１．２％、Ｓｉを０．４〜１．５％、Ｍｏを０．５〜１．５％、Ｃｒを０．５〜２．０％、残部Ｆｅ及び不可避的不純物元素を含有する鉄鋼材料からなり、表面炭素濃度が０．３〜１．３％、表面窒素濃度が０．２〜０．８％となるように、浸炭窒化処理が施されている。また、玉軸受１０は内輪１１と外輪１２との間に転動自在に組み込まれた複数の玉１３と、これらの玉１３を保持する保持器１４とを備えており、内輪１１の外周面には内側軌道溝１５が形成されている。この内側軌道溝１５は外輪１２の内周面に形成された外側軌道溝１６と対向しており、これらの軌道溝１５，１６は幅方向断面が円弧状に形成されている。また図１２は、本発明の一実施形態に係る玉軸受の玉と軌道溝との関係を示す図である。図に示すように、内輪１１の外径面には、その円周方向に沿って所定曲率半径の内側軌道溝１５が形成され、外輪１２の内径面には、その円周方向に沿い且つ上記内側軌道溝１５と径方向で対向した所定曲率半径の外側軌道溝１６が形成され、その内側軌道溝１５と外側軌道溝１６とによって形成される空間内に、複数の玉１３が円周方向に並んで介挿されている。そして、各玉１３は、上記軌道溝１５，１６にそれぞれ転がり接触する。そして、本実施の形態では、上記玉１３の直径をＤａとし、上記外側軌道溝１６の曲率半径をＲｏ、上記内側軌道溝１５の曲率半径をＲｉとした場合に、Ｒｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、Ｒｏ＝０．５１５Ｄａ〜０．５４５Ｄａを満足するように、転動体である玉１３の直径Ｄａや各軌道溝１５，１６の曲率半径Ｒｉ，Ｒｏを決定したものである。
【００４８】
（実施例）
図１１に示した構成に基づく本発明の実施例と比較例を表４に示す。また、表４の比較例Ａ、Ｂおよび実施例Ａ、Ｂの内輪を下記の評価条件で５０００ｍｉｎ^−１〜３１０００ｍｉｎ^−１の速度で回転させたときの外輪の温度変化を図１３に示す。
【００４９】
【表４】

【００５０】
［評価条件］
組合せ　　２列背面組合せ（背面組合せ定位置予圧）
潤滑　　　オイルエア　給油量　０．０３ｃｃ／４ｍｉｎ／１Ｂｒｇ
組込時アキシアルぱね定数　　１００Ｎ／μｍ
（予圧の影響を受けないように同じばね定数に設定）
駆動方法　　ベルト
冷却方法　　外筒冷却　冷却油温室温＋３℃（室温２２℃）、６ＮＬ／ｍｉｎ
評価方法　　温度上昇、焼き付き限界
図１３に示すように、玉径を７．９３８ｍｍ、玉数を２６個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１５°、内外輪材質をＳＵＪ２、玉材質をセラミックとした玉軸受（比較例Ａ）の場合には、内輪の回転速度が２４０００ｍｉｎ^−１を超えると外輪温度が急激に上昇し、内輪の回転速度が２７０００ｍｉｎ^−１に達すると焼付きを起こすことがわかる。これに対し、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝の幅方向断面の曲率半径をＲｉ＝０．５８Ｄａ、外側軌道溝の幅方向断面の曲率半径をＲｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＨＸ（日本精工株式会社商品名、Ｃ；０．２〜１．２重量％、Ｓｉ；０．７〜１．５重量％、Ｍｏ；０．５〜１．５重量％、Ｃｒ；０．５〜２．０重量％、残部Ｆｅ及び不可避的不純物元素を含有する鉄鋼材料、表面炭素濃度：０．８〜１．３％、表面窒素濃度：０．２〜０．８％）、玉材質をセラミックとした玉軸受（比較例Ｂ）の場合には、内輪の回転速度が３１０００ｍｉｎ^−１に達しても焼付きを起さないことがわかる。
【００５１】
また、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＵＪ２、玉材質をセラミックとした玉軸受（実施例Ａ）の場合や、玉径を８．７３１ｍｍ、玉数を２５個、内側軌道溝および外側軌道溝の幅方向断面の曲率半径をＲｉ，Ｒｏ＝０．５４Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１８°、内外輪材質をＳＨＸ、玉材質をセラミックとした玉軸受（実施例Ｂ）の場合も比較例Ｂと同様に、内輪の回転速度が３１０００ｍｉｎ^−１に達しても焼付きを起さないことがわかる。
【００５２】
以上のことから、内側軌道溝の幅方向断面の曲率半径をＲｉ＝０．５３５Ｄａ〜０．５４５Ｄａ、外側軌道溝の幅方向断面の曲率半径をＲｏ＝０．５１５Ｄａ〜０．５４５Ｄａ、内側軌道溝および外側軌道溝に対する玉の接触角を１７°〜２６°、玉の直径を軸受断面高さＨ（外輪の外径から内輪の内径を差し引いた値の１／２）に対して０．４Ｈ〜０．５６Ｈに設定することにより、高速回転に伴う発熱量を抑制して焼付きを防止できると共に転がり疲労寿命を長くすることができ、かつ低コスト化を図ることができる。
【００５３】
【発明の効果】
以上説明したように、本発明によれば、低コストで高速回転での低温度上昇、高剛性、耐荷重性、耐焼き付き性、耐摩耗性に優れた玉軸受を提供できる。
【図面の簡単な説明】
【図１】玉軸受の回転時のアキシアルばね定数を示す図である。
【図２】玉軸受の回転時の内輪ＰＶ値を示す図である。
【図３】玉軸受の回転時の内輪ＰＶ値と外輪ＰＶ値を示す図である。
【図４】玉軸受の発熱量を示す図である。
【図５】玉軸受の転がり疲労寿命を示す図である。
【図６】玉軸受の静的許容アキシアル荷重を示す図である。
【図７】玉軸受の内輪ＰＶ値と接触角との関係を示す図である。
【図８】玉軸受の発熱量と接触角との関係を示す図である。
【図９】玉軸受の外輪ＰＶ値と接触角との関係を示す図である。
【図１０】玉軸受の内輪ＰＶ値と内外輪温度差との関係を示す図である。
【図１１】本発明の一実施形態に係る玉軸受の部分断面図である。
【図１２】本発明の一実施形態に係る玉軸受の玉と軌道溝との関係を示す図である。
【図１３】玉軸受の外輪温度と内輪回転数との関係を示す図である。
【図１４】玉軸受の静的許容アキシアル荷重と内側軌道溝曲率半径との関係を示す図である。
【符号の説明】
１１　　内輪
１２　　外輪
１３　　玉
１４　　保持器
１５　　内側軌道溝
１６　　外側軌道溝[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to a ball bearing used for various machine tools and the like.
[0002]
[Prior art]
2. Description of the Related Art In recent years, various types of machine tools have been accelerated in rotation of a spindle and feed of peripheral devices centering on a machining center in order to improve productivity. Along with this, the ball bearings used as bearings for supporting the main shaft and the rotating shaft of peripheral devices described above are also required to be operated at higher speeds. Suppression, seizure prevention, and high rigidity are required for ball bearings.
[0003]
Conventionally, this type of ball bearing includes an inner ring formed in a ring shape, an outer ring concentrically arranged on the outer periphery of the inner ring, and a plurality of balls rotatably incorporated between the outer ring and the inner ring. And an inner raceway groove is formed on the outer peripheral surface of the inner race. The inner raceway groove is opposed to the outer raceway groove formed on the inner peripheral surface of the outer race, and these raceway grooves have an arc-shaped cross section in the width direction.
[0004]
In such a ball bearing, when the ball rolls on the raceway groove with the rotation of the inner ring or the outer ring, frictional heat is generated between the ball and the raceway groove. The calorific value due to the rotation of such a ball bearing shows a lower value as the product of the contact pressure at the contact surface between the ball and the raceway groove of the inner ring and the outer ring and the sliding speed at the aforementioned contact surface between the ball and the ball are smaller. It is clear that
[0005]
In the above-described conventional ball bearing, the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set to about 0.52 Da when the diameter of the ball is Da, and the bearing cross-section height (the outer diameter of the outer race) If H is １ ／ of the value obtained by subtracting the inner diameter of the inner ring from), the diameter of the ball is set to about Da = 0.57H.
Further, in order to reduce the centrifugal force acting on the ball, a ball bearing whose diameter is set to Da = 0.40H to 0.50H or an inner raceway groove and an outer raceway groove to reduce the gyro moment. A ball bearing having a contact angle of the ball with respect to the angle of 15 ° is commercialized and used.
[0006]
Ball bearings for supporting machine tool spindles are often used at high speeds under a slight amount of oil lubrication. For this reason, oil film rupture is likely to occur, and the contact surface pressure between the ball and the raceway groove may increase due to the effect of the temperature rise of the bearing and the centrifugal force.
On the other hand, due to the above-mentioned high speed, seizure is more likely to occur in the ball bearing. Conventionally, to prevent this seizure, ball bearings have been lubricated with a small amount of oil such as grease lubrication, oil mist lubrication that sprays the lubricating oil onto the bearing, and a small amount of lubricating oil intermittently measured with a fixed amount piston. The lubricating oil is drawn out into the compressed air by the mixing valve, and is changed to oil-air lubrication to be supplied to the bearing as a continuous flow.
[0007]
In addition, high-flow lubrication such as jet lubrication, in which lubricating oil is injected from one or more nozzles into the bearing at a constant pressure, and under-race lubrication, in which lubricating oil is supplied into the bearing from nozzles provided in the inner raceway grooves And the lubrication method has changed. As described above, by changing the lubrication method from a small amount of oil lubrication to a large flow rate lubrication, a sufficient amount of lubricating oil has been supplied into the bearing, and seizing of the bearing has been suppressed to extend the life.
[0008]
In addition, rolling bearings such as ball bearings are sometimes used coaxially side by side. At this time, in order to prevent the relative position in the axial direction of each bearing from changing even during use, a gap between the bearings is conventionally prevented by a fixed position preload using a spacer or the like.
However, when the preload due to the fixed-position preload becomes excessive due to the centrifugal force and the temperature rise due to the high-speed rotation of the bearing, the contact surface pressure between the ball and the raceway groove increases, and seizure easily occurs. Therefore, even if the temperature changes, a constant pressure preload that keeps the load between the bearings substantially constant by using a spring, pneumatic pressure, or hydraulic pressure has been used to suppress the seizure of the bearing and extend the life.
[0009]
Further, a preload switching spindle is also used to secure the rigidity of the ball bearings and to reduce an increase in preload due to a centrifugal force and a temperature rise (effect of a temperature difference between the inner ring and the outer ring) due to high-speed rotation.
In terms of bearing specifications, for example,
{Circle around (1)} In order to suppress the PV value (P; contact surface pressure, V; sliding speed) and heat generation, the radius of curvature of the width direction cross section of the inner raceway groove is set larger than the radius of curvature of the width direction cross section of the outer raceway groove. (See Patent Document 1),
{Circle around (2)} In addition to the above, at least one of the inner ring and the outer ring has a weight ratio of C of 0.2 to 1.2%, Si of 0.7 to 1.5%, and Mo of 0.5 to 1. 5%, 0.5 to 2.0% of Cr, the balance being Fe and steel materials containing unavoidable impurity elements, and a carbonitriding treatment followed by a quenching and tempering treatment to reduce the surface carbon concentration to 0.8 to 2.0%. 1.3% and a low nitrogen material having a surface nitrogen concentration of 0.2 to 0.8% to further reduce heat generation (see Patent Document 2).
(3) In order to suppress the PV value and heat generation, the radius of curvature of the cross section of the inner raceway groove in the width direction is set to be larger than the radius of curvature of the cross section of the outer raceway groove in the width direction. The one that derives optimal parameters of heat generation and PV value (see Patent Document 3),
(4) In order to improve the heat dissipation of the inner ring and reduce the temperature difference from the outer ring, the width of the inner ring is set larger than the width of the outer ring, and the pitch diameter PCD of the ball ((inner ring inner diameter + outer ring outer diameter) / 2) is set smaller than that of a conventional bearing to reduce the rotation and revolving speed of the ball and reduce heat generation of the bearing (see Patent Document 4).
(5) In order to reduce the temperature rise, at least one of the inner ring and the outer ring is formed of powdered high-speed steel having a hardness of HRC65 or more, and the radius of curvature of the raceway groove when formed of the powdered high-speed steel is used as the bearing steel. 1.0 to 1.12 times that of the case formed by the above, the spin slip and gyro slip generated in the Hertz elastic contact surface are reduced, the temperature rise of the bearing is reduced, and the radius of curvature of the raceway groove is increased. (See Patent Document 5)
(6) In order to secure rigidity at low speeds and reduce torque heat generation at high speeds to satisfy both high speed and high rigidity, raceway grooves formed on the outer peripheral surface of the inner ring and the inner peripheral surface of the outer ring One in which the radius of curvature is continuously changed (see Patent Document 6),
(7) In order to reduce the centrifugal force and gyro moment acting on the ball to suppress heat generation and increase the speed, the diameter of the ball is equal to the bearing cross-section height (the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring is 1). / 2) set to 32-35% (see Patent Document 7),
(8) In order to reduce the internal preload and achieve high speed, the outer ring groove bottom thickness is reduced, and the dmn value (dm: ball pitch circle diameter, n: rotation speed) exceeds 2,000,000. In such a case, an outer ring expands and deforms when the internal preload of the bearing increases (see Patent Document 8).
[0010]
[Patent Document 1]
JP-A-11-270564
[Patent Document 2]
Re-published patent WO00 / 37813
[Patent Document 3]
JP 2000-145794 A
[Patent Document 4]
JP-A-10-26132
[Patent Document 5]
Japanese Utility Model Registration No. 2526579
[Patent Document 6]
JP-A-9-177795
[Patent Document 7]
JP-A-10-274244
[Patent Document 8]
JP-A-11-82499
[0011]
[Problems to be solved by the invention]
However, for ball bearings used in machine tool spindles and the like, in addition to demands for high speed and low heat generation, cutting of difficult-to-cut materials has been increasing in recent years, and demands for high speed and high rigidity have been increasing. In addition, the demand for cost reduction is becoming stronger. For this reason, regarding lubrication of ball bearings, oil-air lubrication, oil mist lubrication, and grease lubrication are more desirable than jet lubrication and underrace lubrication, which require a large lubricant supply / discharge device, from the viewpoint of cost reduction.
[0012]
As for the preloading method when a plurality of bearings are juxtaposed along the axis, a fixed position preload that can simplify the preload mechanism is more desirable than a constant pressure preload from the viewpoint of rigidity and cost reduction.
In order to increase the speed and rigidity of ball bearings, if the bearings themselves are highly rigid and resistant to seizure and the amount of heat generated by rotation is low, even if the above-mentioned low-cost lubrication method or preloading method is used, further high-speed Can be achieved.
[0013]
As described in Patent Literature 1 and Patent Literature 2 described above, in order to reduce the heat generation of the bearing and reduce the PV value of the inner ring, the radius of curvature of the widthwise cross section of the inner raceway groove may be increased. However, when actually used for supporting the main shaft of a machine tool, the performance required of a ball bearing not only suppresses heat generation and speeds up, but also includes rigidity and durability life. A load capacity against an excessive static axial load acting on the bearing as an unclamping force when the bearing is removed from the main shaft is also required. Therefore, the radius of curvature of the cross section in the width direction of the inner raceway groove is set to 52.5 to the diameter of the ball. In actually setting the range within a wide range of 60% or less, it is necessary to consider the above-mentioned use conditions for the upper limit of the radius of curvature.
[0014]
The same applies to the case of the ball bearing disclosed in Patent Document 3. The concept of the permissible static axial load of a ball bearing is based on the basic load that the ball rides on the shoulder of either the inner raceway groove or the outer raceway groove and the contact surface pressure when an axial load is applied to the ball bearing. There are two types, the surface pressure limit corresponding to the rated load, and the lower value is usually set as the limit axial load. However, in the case of a running load, 70% of the limit axial load is set as the allowable axial load in consideration of the manufacturing error of the bearing, the variation in the tolerance range, and the safety factor.
[0015]
As for the surface pressure limit, a static allowable load coefficient fs expressed by the following equation is used as a general guide.
fs = Co / Po
Co; Basic static load rating (N)
Po; static equivalent load (N)
Here, when the ball bearing is used for supporting the main shaft of the machine tool, it is desirable to set the static allowable load coefficient fs to 1.5 or more. The sum of the amount and the permanent deformation amount of the raceway is 85% of the maximum stress 4200 MPa (Basic static rated load: Coa because it is a static allowable axial load) that is approximately 0.0001 times the diameter of the rolling element, or 3570 MPa. Since fs can be set to about 1.7 to 1.8, the maximum stress 3570 MPa is set as the static allowable surface pressure.
[0016]
FIG. 14 shows the specifications of the ball bearings as inner ring inner diameter; Di = 70 mm, outer ring outer diameter; Do = 110 mm, ball diameter; Da = 8.731 mm, number of balls; 25, radius of curvature of the outer raceway groove in the width direction; Ro = 0.52 Da, contact angle of the ball with the inner raceway groove and the outer raceway groove; 15 °, inner and outer race material; SUJ2, ball material; Si ₃ N ₄ The relationship between the radius of curvature Ri of the cross section in the width direction of the inner raceway groove and the permissible static axial load in the case of is shown.
[0017]
The use of ceramic balls for high-speed applications is now within the common sense.However, when a load is applied, the modulus of longitudinal elasticity of the ceramic is larger than that of bearing steel. Also, the contact stress between the ball and the raceway groove increases, and the static allowable axial load also decreases.
Static permissible axial loads are limited by permissible stresses prior to riding loads. As shown in FIG. 14, when the radius of curvature of the cross section in the width direction of the inner raceway groove is compared with the conventional ball bearing in which Ri = 0.52 Da is set, when the radius of curvature of the widthwise cross section of the inner raceway groove is increased, the static It can be seen that the ultimate allowable axial load is greatly reduced.
[0018]
When the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is set to 0.58 Da, the static allowable axial load is reduced by 65% as compared with the conventional ball bearing, and the capacity of the machine tool spindle becomes insufficient. Occurs.
When the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is increased, the contact portion of the ball (hereinafter referred to as “contact ellipse”) contacting the raceway groove is reduced during rotation of the bearing. This makes it possible to suppress the sliding of the ball and to reduce the PV value and the heat generation, but naturally the contact ellipse is reduced. For this reason, since the stress of the contact portion becomes larger than that of the conventional ball bearing, the rolling fatigue life naturally decreases.
[0019]
In addition, as important parameters for heat generation, high speed, and rigidity, (1) the radius of curvature of the cross section in the width direction of the width cross section of the raceway groove, (2) the ball diameter, and (3) the contact angle are mentioned. The content does not take the contact angle parameter into consideration in all cases.
In addition, the recent trend of machine tool spindles is that not only high speed but also variation / variable production is performed, so that machining is performed in various rotation ranges, and the frequency of tool conversion is increasing. In order to realize high-efficiency machining under such machining conditions, it is indispensable to be able to rotate at high speed and to shorten the start / stop time (sudden acceleration / deceleration time) of the spindle. During this transition period of rapid rotation fluctuation, changes in the surrounding environment (heat generation of the motor, cooling of the outer cylinder, etc.) are remarkable, and the bearing for the main shaft is exposed to severe thermal load conditions. To put it simply, a temperature difference between the inner and outer rings may occur, and the internal preload of the bearing may increase rapidly to cause seizure. It is necessary to maintain stable performance even under such conditions, and it is necessary to design the bearing insensitively with respect to the PV value.
[0020]
The present invention has been made in view of the above-described problems, and it is an object of the present invention to suppress the amount of heat generated by high-speed rotation, prevent seizure with high rigidity, and increase the rolling fatigue life. It is an object of the present invention to provide a ball bearing capable of securing static axial load capacity when used as a bearing for a machine tool main shaft and reducing cost.
[0021]
[Means for Solving the Problems]
The invention according to claim 1 of the present invention is directed to an inner race having an inner raceway groove on an outer peripheral surface, an outer race having an outer raceway groove facing the inner raceway groove on an inner peripheral surface, the inner raceway groove and the outer raceway groove. And a plurality of balls disposed between the inner raceway groove and the outer raceway groove, wherein the widthwise cross-section of the inner raceway groove and the outer raceway groove are formed in an arc shape, wherein the widthwise cross-section of the inner raceway groove is provided. Is the radius of curvature of Ri, the radius of curvature of the cross section in the width direction of the outer raceway groove is Ro, the diameter of the ball is Da, and 1/2 of the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring is the bearing sectional height. When H, the radius of curvature of the cross section of the inner raceway groove in the width direction is set to Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the cross section of the outer raceway groove in the width direction is Ro = 0.515 Da to 0. .545 Da, and said inner raceway groove and It sets the contact angle of the ball against Kisotogawa raceway grooves in 17 ° ~ 26 °, characterized in that the diameter of the ball is set to Da = 0.4H~0.56H.
[0022]
According to a second aspect of the present invention, in the ball bearing according to the first aspect, the ball is formed of ceramic.
According to a third aspect of the present invention, in the ball bearing according to the second aspect, at least one of the inner ring and the outer ring has a weight ratio C of 0.2 to 1.2% and Si of 0.4 to 0.4. It is made of a steel material containing 1.5%, Mo 0.5 to 1.5%, Cr 0.5 to 2.0%, balance Fe and unavoidable impurity elements, and is subjected to carbonitriding. The surface carbon concentration is set in the range of 0.8 to 1.3%, and the surface nitrogen concentration is set in the range of 0.2 to 0.8%.
[0023]
According to a fourth aspect of the present invention, in the ball bearing according to any one of the first to third aspects, a preload is applied to the ball bearing.
According to a fifth aspect of the present invention, in the ball bearing according to the fourth aspect, the ball bearing is used at dmn of 1,000,000 or more.
According to a sixth aspect of the present invention, in the ball bearing according to the fourth or fifth aspect, the ball bearing is used for a spindle for a machine tool main shaft.
[0024]
Here, the effective range of each of the above-described component elements will be described.
(1) Si: 0.4 to 1.5% by weight
Si is an element having an effect on temper softening resistance, and has an effect of improving high-temperature strength and delaying decomposition of retained austenite, which is effective in preventing indentation-type delamination in a high-temperature environment. If the Si content is less than 0.4% by weight, the high-temperature strength becomes insufficient and the indentation origin type peeling occurs. Therefore, the lower limit is set to 0.4% by weight. On the other hand, if the Si content exceeds 1.5% by weight, the mechanical strength is reduced and carburization is impeded. Therefore, the upper limit is set to 1.5% by weight.
(2) Mo; 0.5 to 1.5% by weight
Mo is an element having an effect on temper softening resistance like Si, and has an effect of improving high-temperature strength. Mo acts as a carbide forming element for forming fine carbides on the carbonitrided surface. If the Mo content is less than 0.5% by weight, the high-temperature strength becomes insufficient and the amount of carbide deposited on the surface becomes insufficient. Therefore, the lower limit is set to 0.5% by weight. On the other hand, if the Mo content exceeds 1.5% by weight, giant carbides are generated at the stage of the material, leading to the loss of carbides and shortening the rolling fatigue life of the bearing. And
(3) Cr: 0.5 to 2.0% by weight
Cr is an additive element having the same function and effect as Mo. If the Cr content is less than 0.5% by weight, the high-temperature strength becomes insufficient and the amount of carbide precipitated on the surface becomes insufficient. Therefore, the lower limit is set to 0.5% by weight. On the other hand, if the Cr content exceeds 2.0% by weight, giant carbides are generated at the stage of the raw material, causing the carbides to fall off and reducing the rolling fatigue life of the bearing. Therefore, the upper limit is set to 2.0% by weight. And
(4) C: 0.2 to 1.2% by weight
As described above, if the amount of retained austenite is too large, the retained austenite is decomposed and the shape changes with time, and the dimensional stability of the bearing is impaired. On the other hand, the presence of retained austenite on the inner ring surface and the outer ring surface is effective in preventing indentation origin type peeling. Therefore, it is necessary to suppress the amount of retained austenite in the entire bearing in a state where the retained austenite is present on the surface. From such a viewpoint, it is preferable that the amount of the average retained austenite in the steel including the surface and the core is 5% by volume or less, and for that purpose, the carbon concentration on which the retained austenite depends is reduced to 1.2% by weight or less. Therefore, the upper limit was set to 1.2% by weight. On the other hand, if the carbon concentration is less than 0.2% by weight, it takes a long time to obtain a desired carburizing depth by the carbonitriding treatment, which causes an increase in the overall cost. 0.2% by weight.
(5) Surface carbon concentration: 0.8 to 1.3% by weight
When carbon is added to the surface by carbonitriding, the martensitic structure serving as the matrix can be solid-solution strengthened, and a large amount of retained austenite effective in preventing indentation origin type exfoliation can be formed in the very surface layer. If the surface carbon concentration is less than 0.8% by weight, the surface hardness becomes insufficient and the rolling fatigue life and wear resistance are reduced. Therefore, the lower limit is set to 0.8% by weight. On the other hand, if the surface carbon concentration exceeds 1.3% by weight, giant carbides precipitate during carbonitriding treatment and the rolling fatigue life is reduced. Therefore, the upper limit was set to 1.3% by weight.
(6) Surface nitrogen concentration: 0.2 to 0.8% by weight
When nitrogen is added to the surface by carbonitriding, the tempering resistance is improved, the high-temperature strength is increased, the wear resistance is improved, and a large amount of retained austenite is present in the very surface layer, which is effective in preventing indentation type delamination. Can be. If the surface nitrogen concentration is less than 0.2% by weight, the high-temperature strength is reduced and the wear resistance is improved, so the lower limit was set to 0.2% by weight. On the other hand, if the surface nitrogen concentration exceeds 0.8% by weight, it becomes difficult to finish the grinding during the production of the bearing, and the productivity of the bearing is reduced due to difficult grinding. .
(7) Other component elements
It is preferable to add a small amount of Ti as another component element. This is because, when Ti is added, fine titanium carbide (TiC) or carbonitride (Ti (C + N)) precipitates and disperses in the matrix and improves wear resistance and seizure resistance. In this case, the Ti content is desirably 0.1 to 0.3% by weight. If the Ti content is less than 0.1% by weight, the effect of carbide precipitation cannot be obtained, so the lower limit is set to 0.1% by weight. On the other hand, if the Ti content exceeds 0.3% by weight, a huge precipitate is likely to be formed, and this may become a defect to reduce the rolling fatigue life, so that the upper limit is set to 0.3% by weight. %. Incidentally, when the size of the titanium precipitate (TiC, Ti (C + N)) is 0.1 μm or less, it contributes to improvement of wear resistance and seizure resistance.
[0025]
BEST MODE FOR CARRYING OUT THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings. Before describing the embodiments of the present invention, the circumstances that led to the invention of the ball bearing according to the present invention will be described.
In order to find the specifications of the most suitable ball bearing for supporting the main shaft of a machine tool, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, the ball diameter to 8.731 mm, the number of balls to 25, and the contact angle to 25. 18 °, inner and outer ring material is SUJ2, ball material is Si ₃ N ₄ Axial rigidity, inner ring PV value, outer ring PV value, heat value when the radius of curvature Ri, Ro of the width direction cross section of the inner raceway groove and the outer raceway groove are combined with the ball diameter Da as shown in Table 1. , Rolling fatigue life and static allowable axial load were calculated. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , Back combination (spacing width 20 mm), shaft fitting; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (clearance fit, loose fit amount: 8 μm), lubrication; oil-air lubrication, assembly Preload load when axial spring constant becomes 100 N / μm (Axial spring constant 100 N / μm when this is incorporated is used as a measure of rigidity that can be machined), temperature condition: outer ring temperature rise 24 ° C., inner and outer ring temperature difference 16 Calculated under the condition of ° C.
[0026]
[Table 1]

[0027]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 FIG. 1 shows the axial spring constant of (1). As shown in the figure, the ball bearing having the highest axial spring constant during rotation among the numbers 1 to 25 is the ball bearing of number 1, and the ball bearing having the lowest axial spring constant during rotation is number 25. Ball bearings.
Comparing the ball bearing of No. 1 with the ball bearing of No. 25, the ball bearing of No. 1 has a radius of curvature Ri, Ro of 0.52 Da in the width direction cross section of the inner raceway groove and the outer raceway groove when the diameter of the ball is Da. On the other hand, in the ball bearing of No. 25, since the curvature radii Ri and Ro of the width direction cross sections of the inner raceway groove and the outer raceway groove are 0.60 Da, the widthwise cross-sections of the inner raceway groove and the outer raceway groove are 0.65 Da. It can be seen that the larger the radii of curvature Ri and Ro, the lower the axial spring constant during rotation and the lower the axial rigidity during rotation.
[0028]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 ) Are shown in FIG. As shown in the figure, the ball bearing with the highest inner ring PV value during rotation among the numbers 1 to 25 is the ball bearing with the number 1 and the ball bearing with the lowest inner ring PV value during the rotation is number 20. And 25 ball bearings, as well as ball bearings with numbers 5, 10 and 15.
[0029]
In the ball bearings of Nos. 5, 10, 15, 20, and 25, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is 0.60 Da, assuming that the diameter of the ball is Da. It can be seen that the inner ring PV value does not change if the radius of curvature Ri of the width direction cross section of the inner raceway groove is the same value even if the radius of curvature Ro of the inner raceway changes. It can be seen that the inner ring PV value can be kept low.
[0030]
The inner ring of each ball bearing indicated by the numbers 1 to 25 in Table 1 is 25000 min. ^-1 FIG. 3 shows the inner wheel PV value and the outer wheel PV value when rotating at. As shown in FIG. ^-1 , The inner wheel PV value is larger than the outer wheel PV value. The reason for this is that at high speeds, the outer ring is controlled (the outer ring and the ball are close to pure rolling), and the rotation axis of the ball is based on the contact angle on the outer ring side. Is increased. Therefore, it is desirable to use the inner ring PV value as an index for determining image sticking.
[0031]
When each of the ball bearings indicated by numbers 1 to 25 in Table 1 rotates (25,000 min. ^-1 4) shows the calorific value. As shown in the figure, the ball bearing with the highest heat value among the numbers 1 to 25 is the ball bearing with the number 1 and the ball bearing with the lowest heat value with the ball bearings with the numbers 10, 15, 20, and 25. It is a bearing. In the ball bearings of Nos. 10, 15, 20, and 25, the radius of curvature Ri of the cross-section in the width direction of the inner raceway groove is Ri = 0.60 Da. Therefore, the larger the radius of curvature Ri of the cross-section of the inner raceway groove in the width direction is, the larger the radius. It can be seen that the amount of heat generated during high-speed rotation is reduced, and that seizure can be suppressed.
[0032]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 5) shows the rolling fatigue life. As shown in the figure, the ball bearing with the longest rolling fatigue life among the numbers 1 to 25 is the ball bearing with the number 6 and the ball bearing with the shortest rolling fatigue life is the ball bearing with the number 25.
Comparing the ball bearing No. 6 with the ball bearing No. 25, the ball bearing No. 6 has a radius of curvature Ri = 0.52 Da in the width direction cross section of the inner raceway groove, and a radius of curvature in the width direction cross section of the outer raceway groove. While Ro is Ro = 0.54 Da, the ball bearing of No. 25 has a radius of curvature Ri = 0.60 Da in the width direction cross section of the inner raceway groove, and the radius of curvature Ro of the width direction cross section of the outer raceway groove is Ro. = 0.60 Da, it can be seen that the larger the radius of curvature Ri, Ro of the cross section in the width direction of the inner raceway groove and the outer raceway groove, the lower the rolling fatigue life.
[0033]
At the time of rotation of each ball bearing indicated by number 1 to 25 in Table 1 (25000 min ^-1 6) shows the permissible static axial load of FIG. As shown in the drawing, the ball bearing having the highest static allowable axial load among the numbers 1 to 25 is the ball bearing of the number 1, and the ball bearings having the lowest static allowable axial load are the numbers 5 and 10 on the contrary. , 15, 20, and 25 ball bearings.
[0034]
In the ball bearings of Nos. 5, 10, 15, 20, and 25, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is Ri = 0.60 Da. Therefore, the radius of curvature Ri of the cross section in the widthwise direction of the inner raceway groove is smaller. It is understood that the larger the value, the lower the allowable static axial load.
Therefore, in order to reduce the calorific value and the PV value of the ball bearing, the radius of curvature of the widthwise cross section of the inner raceway groove and the outer raceway groove may be increased, but the widthwise cross section of the inner raceway groove and the outer raceway groove may be reduced. When the radius of curvature is increased, the rigidity and life of the bearing are reduced, and the allowable static load is reduced. Conversely, if the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is reduced, the calorific value and the PV value increase.
[0035]
For machine tool spindle support, a well-balanced bearing is required for all of the seizure index, calorific value, axial rigidity, rolling fatigue life, and static allowable axial load. Compared the calculation results of the axial spring constant, the inner ring PV value, the calorific value, the rolling fatigue life, and the static allowable axial load of each ball bearing indicated by numbers 1 to 25 by absolute value evaluation. In the absolute value evaluation, the best value was 100 for each condition and the worst value was 0 for each condition, and the absolute value was evaluated by indexing the difference. The burn-in index—the inner-ring PV value, which is the most important under high-speed conditions, was increased in importance and calculated with the rank being doubled.
[0036]
The calculation formula used in the above absolute value evaluation is shown below.
(1) Rigidity (Ka)
KaRank = 100− (Ka _Max −Ka) / ((Ka _Max -Ka _Min ) / 100)
Ka _Max The axial spring constant showing the maximum value in FIG.
Ka _Min The axial spring constant showing the minimum value in FIG.
(2) Inner ring PV value (Pv)
PvRank = (Pv _Max −Pv) / ((Pv _Max -Pv _Min ) / 100)
Pv _Max The inner ring PV value showing the maximum value in FIG.
Pv _Min The PV of the inner ring showing the minimum value in FIG.
(3) Calorific value (W)
WRank = (W _Max −W) / ((W _Max -W _Min ) / 100)
W _Max The calorific value showing the maximum value in FIG.
W _Min The calorific value showing the minimum value in FIG.
(4) Rolling fatigue life (L)
LRank = 100− (L _Max −L) / ((L _Max -L _Min ) / 100)
L _Max A rolling fatigue life showing the maximum value in FIG.
L _Min Rolling fatigue life showing the minimum value in FIG.
(5) Allowable static load (Fa)
FaRank = 100− (Fa _Max −Fa) / ((Fa _Max -Fa _Min ) / 100)
Fa _Max A static allowable axial load showing the maximum value in FIG.
Fa _Min A static allowable axial load showing the minimum value in FIG.
(6) TOTAL-Rank
KaRank + (PvRank × 2) + WRank + LRank + FaRank
Table 2 shows the calculation results of the above-described KaRank, PvRank, WRank, LRank, FaRank, and TOTAL-Rank.
[0037]
[Table 2]

[0038]
Regarding the evaluation points of the TOTAL-Rank, the ball bearings of Nos. 2 and 7 are the most excellent, and can be said to be the optimum specifications. The TOTAL-Rank of the numbers 2 and 7 has almost the same value.
In the bearing specification in which any index takes 100 points from the Rank coefficient in Table 2, any value is 0 or a value close to 0, which is the extreme. For this reason, it cannot be said that it is a very good specification for a machine tool spindle.
[0039]
From the calculation results shown in Table 2, the ball bearing most suitable for supporting the main shaft of the machine tool has a radius of curvature Ri of the widthwise cross section of the inner raceway groove of Ri = 0.54 Da, where Da is the diameter of the ball, and It can be seen that the radius of curvature Ro of the cross section in the width direction of the outer raceway groove is Ro = 0.52 Da or 0.54 Da. However, when the bearing is actually manufactured, a manufacturing error is also included. Therefore, the radius of curvature Ri of the cross section in the width direction of the inner raceway groove is Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the cross section in the width direction of the outer raceway groove is Ri. It is desirable to set Ro to be in a range of Ro = 0.515 Da to 0.545 Da.
[0040]
Next, to confirm the influence of the contact angle of the balls, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, the ball diameter to Da = 8.731 mm, the number of balls to 25, the inner raceway groove. The curvature radius of the cross section of the outer raceway groove in the width direction is Ri = Ro = 0.54 Da, the inner and outer race materials are SUJ2, and the ball material is Si. ₃ N ₄ And the inner ring PV value, the outer ring PV value, and the calorific value were calculated when the contact angles of the balls were α = 12 °, 15 °, 18 °, 20 °, 22 °, 25 °, and 30 °. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , DB combination (spacing width 20 mm), shaft fitting; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (gap fit, loose fit amount: 8 μm), lubrication; oil-air lubrication, assembly Preload load when axial spring constant becomes 100 N / μm (Axial spring constant 100 N / μm when this is incorporated is used as a measure of rigidity that can be machined), temperature condition: outer ring temperature rise 24 ° C., inner and outer ring temperature difference 16 Calculated under the condition of ° C. 7 to 9 show the calculation results.
[0041]
As shown in FIG. 7, when the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is α = 30 °, the slip of the ball increases due to the effect of the gyro moment, and the inner ring PV value sharply increases. . Therefore, in order to keep the inner ring PV value low, the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove needs to be smaller than 30 °. For example, when the angle is set to 12 °, the axial rigidity decreases, and when the axial spring constant at the time of installation is set to 100 N, an excessive preload must be applied, and the preload during rotation increases. For this reason, as shown in FIG. 7, since the inner ring PV value is larger than that of a ball bearing in which the contact angle of the ball with the inner raceway groove and the outer raceway groove is 15 °, the widthwise cross section of the inner raceway groove and the outer raceway groove is increased. Is set to Ri, Ro = 0.54 Da as shown in Table 1, the contact angle of the ball to the inner raceway groove and the outer raceway groove is preferably set to 15 ° to 25 °. .
[0042]
Further, as shown in FIG. 8, when the contact angle of the ball with the inner raceway groove and the outer raceway groove is α = 30 °, the ball slides more due to the effect of the gyro moment, and the heat generation amount rises sharply. I do. Therefore, in order to keep the calorific value low, the contact angle of the ball with the inner raceway groove and the outer raceway groove must be smaller than 30 °, but the contact angle of the ball with the inner raceway groove and the outer raceway groove is 15 °. In the case of °, the calorific value increases more than when the contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °. For this reason, when the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set as Ri, Ro = 0.54 Da as shown in Table 1, the ball for the inner raceway groove and the outer raceway groove is Is desirably 18 ° to 25 °.
[0043]
Further, as shown in FIG. 9, when the contact angle of the ball with the inner raceway groove and the outer raceway groove is α = 15 ° or less, the outer ring PV value increases for the same reason as described above. Therefore, in order to keep the outer ring PV value low, the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove needs to be larger than 15 °. In the case of 30 °, the sliding of the ball becomes large due to the influence of the gyro moment, and the outer ring PV value increases as compared with the case where the contact angle of the ball with the inner raceway groove and the outer raceway groove is 25 °. For this reason, when the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is set to Ri, Ro = 0.54a as shown in Table 7, the ball for the inner raceway groove and the outer raceway groove is Is desirably 18 ° to 25 °.
[0044]
As described above, from the results of FIGS. 7 to 9, it can be said that the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is desirably set to 18 ° to 25 °. However, when the bearing is actually manufactured, a manufacturing error is included. Therefore, it is desirable that the contact angle of the ball with respect to the inner raceway groove and the outer raceway groove is 17 ° to 26 °.
Next, in order to confirm the ball diameter, the present inventors set the inner ring inner diameter to 70 mm, the outer ring outer diameter to 110 mm, and set the curvature radius of the inner raceway groove and the outer raceway groove in the width direction cross section to Ri = Ro = 0.54 Da. The contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, the material of the inner and outer rings is SUJ2, and the material of the ball is Si. ₃ N ₄ And the ball diameter, the number of balls, and the cross-sectional ratio (= ball diameter / (outer ring outer diameter−inner ring inner diameter) / 2) × 100) were calculated as the combinations shown in Table 3 to calculate the inner ring PV value. Note that the calculation here is performed at a rotation speed of 25000 min. ^-1 , DB combination (spacing width 20 mm), shaft fit; T21 (tight fit, tight fit amount: 21 μm), housing fit; L8 (gap fit, loose fit amount: 8 μm), lubrication; oil-air lubrication, assembly Calculated under the conditions of an axial spring constant of 100 N / μm (the axial spring constant of 100 N / μm at the time of assembling is used as a measure of rigidity that can be machined), temperature conditions, and an outer ring temperature rise of 24 ° C. FIG. 10 shows the calculation results.
[0045]
[Table 3]

[0046]
As is clear from FIG. 10, when the diameter of the ball is 7.144 mm and 7.938 mm, when the temperature difference between the inner ring and the outer ring increases, the inner ring PV value sharply increases, and the temperature between the inner ring and the outer ring increases. When the difference is around 30 °, the inner ring PV value is about 1900 to 2100 MPa · m / s. When the diameters of the balls are 11.906 mm and 12.7 mm, the change in the inner ring PV value is gradual, but the original PV value itself is high, and the temperature difference between the inner ring and the outer ring is about 30 °. The inner ring PV value is about 1900 to 2100 MPa · m / s as in 7.144 mm and 7.938 mm. On the other hand, when the diameter of the ball is 8.731 mm, 9.525 mm, 10.319 mm, and 11.112 mm, the inner ring PV value increases as the temperature difference between the inner ring and the outer ring increases. When the temperature difference with the inner ring is around 30 °, the inner ring PV value is about 1800 MPa · m / s. Therefore, if the bearing cross-section height (1/2 of the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring) is H, if the diameter Da of the ball is from 0.437H to 0.556H, the inner ring is Since the rate of change of the inner ring PV value with respect to the temperature difference between the outer ring and the outer ring is substantially the same, the temperature difference between the inner ring and the outer ring can be reduced by setting the diameter Da of the ball to 0.437H to 0.556H. However, the PV value can be reduced.
[0047]
FIG. 11 is a partial sectional view of a ball bearing according to one embodiment of the present invention. As shown in FIG. 1, a ball bearing 10 according to an embodiment of the present invention includes an inner ring 11 formed in a ring shape, and an outer ring 12 arranged concentrically around the outer periphery of the inner ring 11. At least one of the inner ring 11 and the outer ring 12 has a weight ratio of C of 0.2 to 1.2%, Si of 0.4 to 1.5%, Mo of 0.5 to 1.5% and Cr of 0%. It is made of a steel material containing 0.5 to 2.0%, balance Fe and unavoidable impurity elements, and has a surface carbon concentration of 0.3 to 1.3% and a surface nitrogen concentration of 0.2 to 0.8%. Thus, the carbonitriding process is performed. Further, the ball bearing 10 includes a plurality of balls 13 rotatably incorporated between the inner ring 11 and the outer ring 12 and a retainer 14 for holding these balls 13. Has an inner raceway groove 15 formed therein. The inner raceway groove 15 is opposed to an outer raceway groove 16 formed on the inner peripheral surface of the outer race 12, and the raceway grooves 15, 16 are formed to have a circular cross section in the width direction. FIG. 12 is a view showing a relationship between a ball and a raceway groove of a ball bearing according to an embodiment of the present invention. As shown in the drawing, an inner raceway groove 15 having a predetermined radius of curvature is formed on the outer diameter surface of the inner race 11 along the circumferential direction, and the inner raceway surface of the outer race 12 is formed along the circumferential direction on the inner diameter surface of the outer race 12. An outer raceway groove 16 having a predetermined radius of curvature radially opposed to the inner raceway groove 15 is formed, and a plurality of balls 13 are circumferentially arranged in a space formed by the inner raceway groove 15 and the outer raceway groove 16. They are inserted side by side. Then, each ball 13 comes into rolling contact with the track grooves 15 and 16 respectively. In the present embodiment, when the diameter of the ball 13 is Da, the radius of curvature of the outer raceway groove 16 is Ro, and the radius of curvature of the inner raceway groove 15 is Ri, Ri = 0.535 Da-0. The diameter Da of the ball 13 as a rolling element and the radii of curvature Ri and Ro of the raceway grooves 15 and 16 are determined so as to satisfy 0.545 Da and Ro = 0.515 Da to 0.545 Da.
[0048]
(Example)
Table 4 shows examples of the present invention and comparative examples based on the configuration shown in FIG. The inner rings of Comparative Examples A and B and Examples A and B in Table 4 were subjected to the following evaluation conditions for 5000 minutes. ^-1 ~ 31000min ^-1 FIG. 13 shows the temperature change of the outer ring when the rotor is rotated at the speed shown in FIG.
[0049]
[Table 4]

[0050]
[Evaluation conditions]
Combination 2-row back-to-back combination (back-to-back combination fixed position preload)
Lubrication Oil air Oil supply amount 0.03cc / 4min / 1Brg
Axial screw constant at assembly 100N / μm
(The same spring constant is set so as not to be affected by preload.)
Drive method Belt
Cooling method Outer cylinder cooling Cooling oil temperature room temperature + 3 ° C (room temperature 22 ° C), 6NL / min
Evaluation method Temperature rise, burn-in limit
As shown in FIG. 13, the ball diameter is 7.938 mm, the number of balls is 26, the radius of curvature of the cross-section in the width direction of the inner raceway groove and the outer raceway groove is Ri, Ro = 0.54 Da, the inner raceway groove and the outer raceway groove. In the case of a ball bearing (comparative example A) in which the contact angle of the ball with respect to the ball is 15 °, the material of the inner and outer rings is SUJ2, and the material of the ball is ceramic (Comparative Example A), the rotation speed of the inner ring is 24000 min. ^-1 , The outer ring temperature rises sharply, and the rotation speed of the inner ring becomes 27000 min. ^-1 It can be seen that seizure occurs when the temperature reaches. On the other hand, the ball diameter is 8.731 mm, the number of balls is 25, the curvature radius of the inner raceway groove in the width direction cross section is Ri = 0.58 Da, and the curvature radius of the outer raceway groove in the width direction cross section is Ro = 0.54 Da. The contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, and the material of the inner and outer races is SHX (trade name of Nippon Seiko Co., Ltd., C: 0.2 to 1.2% by weight, Si; 0.7 to 1. 5% by weight, Mo: 0.5 to 1.5% by weight, Cr: 0.5 to 2.0% by weight, steel material containing balance Fe and inevitable impurity elements, surface carbon concentration: 0.8 to 1 0.3%, surface nitrogen concentration: 0.2 to 0.8%), and in the case of a ball bearing using ceramic as a ball material (Comparative Example B), the rotation speed of the inner ring was 31,000 min. ^-1 It can be seen that no seizure occurs even when the temperature reaches.
[0051]
Further, the ball diameter is 8.731 mm, the number of balls is 25, the radius of curvature of the width direction cross section of the inner raceway groove and the outer raceway groove is Ri, Ro = 0.54 Da, and the contact angle of the ball with the inner raceway groove and the outer raceway groove. 18 °, ball bearings made of SUJ2 inner and outer ring materials and ceramic ball material (Example A), or a ball diameter of 8.731 mm, 25 balls, width direction of inner raceway groove and outer raceway groove In the case of a ball bearing in which the radius of curvature of the cross section is Ri, Ro = 0.54 Da, the contact angle of the ball with the inner raceway groove and the outer raceway groove is 18 °, the material of the inner and outer rings is SHX, and the material of the ball is ceramic (Example B). In the same manner as in Comparative Example B, the rotation speed of the inner ring was 31,000 min. ^-1 It can be seen that no seizure occurs even when the temperature reaches.
[0052]
From the above, the curvature radius of the cross section in the width direction of the inner raceway groove is Ri = 0.535 Da to 0.545 Da, the curvature radius of the cross section in the width direction of the outer raceway groove is Ro = 0.515 Da to 0.545 Da, and the inner raceway groove is The contact angle of the ball with the outer raceway groove is 17 ° to 26 °, and the diameter of the ball is 0.4H or more with respect to the bearing cross-sectional height H (1/2 of the value obtained by subtracting the inner diameter of the inner ring from the outer diameter of the outer ring). By setting it to 0.56H, the amount of heat generated due to high-speed rotation can be suppressed, seizure can be prevented, the rolling fatigue life can be prolonged, and the cost can be reduced.
[0053]
【The invention's effect】
As described above, according to the present invention, it is possible to provide a ball bearing excellent in low cost, low temperature rise at high speed rotation, high rigidity, load resistance, seizure resistance, and wear resistance at low cost.
[Brief description of the drawings]
FIG. 1 is a view showing an axial spring constant when a ball bearing rotates.
FIG. 2 is a diagram showing an inner ring PV value during rotation of a ball bearing.
FIG. 3 is a diagram showing an inner ring PV value and an outer ring PV value during rotation of a ball bearing.
FIG. 4 is a diagram showing a calorific value of a ball bearing.
FIG. 5 is a view showing a rolling fatigue life of a ball bearing.
FIG. 6 is a diagram showing a static allowable axial load of a ball bearing.
FIG. 7 is a diagram showing a relationship between an inner ring PV value of a ball bearing and a contact angle.
FIG. 8 is a diagram showing a relationship between a calorific value of a ball bearing and a contact angle.
FIG. 9 is a diagram showing a relationship between an outer ring PV value of a ball bearing and a contact angle.
FIG. 10 is a view showing a relationship between an inner ring PV value of a ball bearing and an inner and outer ring temperature difference.
FIG. 11 is a partial sectional view of a ball bearing according to one embodiment of the present invention.
FIG. 12 is a view showing a relationship between a ball and a raceway groove of a ball bearing according to one embodiment of the present invention.
FIG. 13 is a diagram showing the relationship between the temperature of the outer race of the ball bearing and the rotation speed of the inner race.
FIG. 14 is a view showing a relationship between a static allowable axial load of a ball bearing and a radius of curvature of an inner raceway groove.
[Explanation of symbols]
11 Inner ring
12 Outer ring
13 balls
14 Cage
15 inner raceway groove
16 Outer raceway groove

Claims (6)

An inner ring having an inner raceway groove on an outer peripheral surface, an outer race having an outer raceway groove facing the inner raceway groove on an inner peripheral surface, and a plurality of balls disposed between the inner raceway groove and the outer raceway groove; A ball bearing in which a cross section in the width direction of the inner raceway groove and the outer raceway groove is formed in an arc shape,
The radius of curvature of the cross section of the inner raceway in the width direction is Ri, the radius of curvature of the cross section of the outer raceway in the width direction is Ro, the diameter of the ball is Da, and the value obtained by subtracting the inner diameter of the inner race from the outer diameter of the outer race. When 1/2 is the bearing section height H, the radius of curvature of the width section of the inner raceway groove is set to Ri = 0.535 Da to 0.545 Da, and the radius of curvature of the width section of the outer raceway groove is Ri. Is set to 0.515 Da to 0.545 Da, the contact angle of the ball with the inner raceway groove and the outer raceway groove is set to 17 ° to 26 °, and the diameter of the ball is set to Da = 0. A ball bearing characterized by being set to 4H to 0.56H. The ball bearing according to claim 1, wherein the ball is formed of ceramic. At least one of the inner ring and the outer ring has a weight ratio C of 0.2 to 1.2%, Si of 0.4 to 1.5%, Mo of 0.5 to 1.5%, and Cr of 0.1 to 1.5%. It is made of a steel material containing 5 to 2.0%, the balance being Fe and unavoidable impurity elements, and is subjected to carbonitriding, so that the surface carbon concentration is set in the range of 0.8 to 1.3%. 3. The ball bearing according to claim 2, wherein the surface nitrogen concentration is set in a range of 0.2 to 0.8%. The ball bearing according to any one of claims 1 to 3, wherein the ball bearing is used with a preload applied thereto. 5. The ball bearing according to claim 4, wherein the ball bearing is used at a dmn of 1,000,000 or more. The ball bearing according to claim 4, which is used for a spindle for a machine tool main shaft.

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